Safety brake system for trailers

ABSTRACT

In a gaseous-fuelled stoichiometric compression ignition internal combustion engine, a pilot fuel is injected directly into the combustion chamber to help initiate a multi-point ignition. The engine provides performance improvements approaching those of high pressure direct injection engines but with less complexity because the gaseous fuel is introduced into the intake air subsystem at relatively low pressure and as a result of the stoichiometric combustion, the low oxygen content in the combustion products exiting the combustion chamber allows the use of a three-way catalyst instead of other after treatment arrangements normally associated with conventional compression ignition engines that require the addition of a reductant.

FIELD OF THE INVENTION

The present invention relates to a gaseous-fuelled stoichiometriccompression ignition internal combustion engine and, in particular, toengines operating with pilot ignition.

BACKGROUND OF THE INVENTION

Internal combustion engines have been used to produce power and drivemachines for over a century. From the beginning, internal combustionengines have undergone many improvements to become more efficient, morepowerful, and/or less polluting. Various modifications to engine designalong with many alternative fuel choices have been considered. In thisdisclosure, gaseous fuels are fuels that are in the gaseous phase atatmospheric pressure and temperature, and are gases that are combustiblein an internal combustion engine of the disclosed type, with examples ofsuch gaseous fuels being methane, ethane, propane, and other lighterflammable hydrocarbon derivatives as well as hydrogen and natural gasand other mixtures thereof. In particular, natural gas, being cleanerburning relative to conventional diesel fuels, and being abundant andmore widely distributed around the world, has been receiving renewedattention as a substitute for more traditional fuels such as gasolineand diesel. That is, factors such as price, availability, energysecurity, and environmental concerns are leading more fuel users toconsider alternative fuel options.

Natural gas has been used as a fuel in vehicular internal combustionengines for over fifty years. Historically, natural gas driven vehicleswere naturally fumigated, meaning that natural gas was introduced intothe intake manifold, with a mixture of fuel and intake air fed into thecylinders through the open intake valve. With such engines, the mostcommon approach for igniting a gaseous-fuel/air mixture is to employspark ignition because unlike liquid fuels like diesel, gaseous fuelsare generally more difficult to ignite by compression ignition.

Generally there are two types of spark ignited gaseous-fuelled enginesthat have been commercialized, namely so-called lean burn engines thatdeliver an excess amount of oxygen to the combustion chamber, andengines that operate in a stoichiometric mode in which thegaseous-fuel/air mixture is controlled so that during combustionessentially all of the fuel is combined with essentially all of the freeoxygen. That is, with an ideal stoichiometric fuel/air mixture there isjust enough oxygen to burn essentially all of the available fuel. Leanburn engines and stoichiometric engines each have their advantages anddisadvantages. For example, lean burn engines generally allow highercompression ratios and combined with lower throttling losses, this canprovide higher efficiency and lower fuel consumption. A disadvantage oflean burn engines is that the presence of excess oxygen in the exhaustgas exiting the combustion chamber makes a lean burn engine incompatiblewith modern three way catalyst aftertreatment subsystems, which meansthat a more expensive aftertreatment subsystem is required to reduceNO_(x) levels.

Stoichiometric engines normally have lower compression ratios comparedto lean burn engines, which normally results in lower efficiency and/orlower power output, but the combustion products are compatible withmodern three-way catalyst aftertreatment subsystems so this has helpedstoichiometric engines to meet recent emissions standards withoutrequiring the more complex and more expensive aftertreatment subsystemsneeded by lean burn engines. For example, the applicant's relatedcompany, Cummins Westport Inc. recently offered an advanced natural gasengine that operates in a stoichiometric mode, with exhaust gasrecirculation and spark ignition, and engines with this combination offeatures are referred to herein as SESI engines. Compared to earlierengines, it uses relatively high rates of cooled exhaust gasrecirculation (EGR) to reduce excess air and thereby reduce theproduction of NO_(x) during combustion, while also lessening thelikelihood of combustion knock.

Another approach for natural gas engines is not stoichiometric andinvolves the use of compression ignition to ignite the fuel/air mixture(the diesel principle) instead of spark ignition. Higher compressionratios are used than those used in spark ignited engines, thus allowingfor greater power and efficiency. However, as noted previously, a chargeconsisting of gaseous fuel and air is difficult to ignite by compressionalone without the use of an ignition assisting device, such as theignition of a more readily ignited pilot fuel, such as a small amount ofdiesel fuel, or a glow plug or other hot surface.

When a pilot fuel is used it is typically directly injected into thecombustion chamber of the engine cylinders to initiate ignition of theprimary gaseous fuel. The pilot fuel mixes with air in the combustionchamber, ignites as a result of the pressure/temperature conditionstherein, and in turn ignites the gaseous fuel. The amount of pilot fuelrequired can be very small, for instance approximately 1% of the totalfuel present. Such pilot operation is sometimes referred to as“micropilot” and this term is defined herein to mean this.

Engines using a compression ignition approach and operating primarilywith fumigated gaseous fuel are often referred to as “dual fuel” enginesand are referred to herein as such. Dual fuel engines can inject dieselpilot fuel directly into the combustion chamber for ignition purposesand EGR can be employed. However, this approach uses an excess amount ofair since it does not employ a throttle and therefore it is notstoichiometric, and like lean burn engines, dual fuel engines requiremore complicated and expensive exhaust treatment to treat emissions. Anadvantage of dual fuel engines is that they allow for a relatively easyretrofit of existing diesel engines. In addition, it allows for the useof diesel only (100% pilot fuel) should that prove desirable ornecessary.

In general, engines can be made more efficient, more powerful, and lesspolluting with more precise control over the timing for fuel injection,the quantity of fuel injected, and the rate of fuel injection during aninjection event. Better efficiency and emissions can be achieved in agaseous-fuelled engine if the gaseous fuel is injected directly into thecylinders under high pressure with the timing for start of injectionbeginning near the end of the compression stroke of the piston. Thisapproach reduces the potential for combustion knock and allowsgaseous-fuelled engines to be operated with the same compression ratiosas conventional diesel engines. However, this requires a morecomplicated and expensive fuel supply subsystem which can deliver boththe primary gaseous fuel and the pilot fuel at injection pressures of atleast 200 bar.

Advanced engines using direct injection of gaseous fuels into thecombustion chambers of the engine cylinders at such injection pressuresare disclosed, for example, in co-owned U.S. Pat. Nos. 6,073,862,6,439,192 and 6,761,325. Therein and herein, these engines are referredto as high pressure direct injection engines or “HPDI engines”. Whileoffering advantages compared to other gaseous-fuelled engines in termsof power, efficiency and high potential substitution percentages ofprimary gaseous fuel for diesel, such engines operate in a lean mode,with excess air (not stoichiometric), like conventional diesel engines.Accordingly, to comply with current emissions requirements in manyjurisdictions, compared to stoichiometric engines, HPDI enginestypically require a more complicated and expensive aftertreatmentsubsystem for treatment of the exhaust.

A variation of HPDI uses a glow plug or other hot surface ignitiondevice instead of a pilot fuel, to ignite the gaseous fuel. Engines thatuse this approach are disclosed, for example, in co-owned U.S. Pat. Nos.6,845,746, 7,077,115 and 7,281,514. In the disclosed preferredembodiments, a gaseous fuel is injected directly into the combustionchamber, with the timing for start of injection being late in thecompression cycle near or at top dead center and at about the sameinjection pressure as HPDI engines that employ a pilot fuel.

Numerous other engine embodiments have been contemplated and disclosedin the art where the primary fuel is other than natural gas. Forinstance, the Southwest Research Institute (SWRI), in U.S. Pat. No.6,679,224, discloses a diesel engine employing EGR that is adapted towork temporarily under stoichiometric conditions, and in particular toprovide a means for regenerating a lean NO_(x) trap without introducingunburned fuel into the exhaust stream of the engine, or requiringadditional substances for operating the engine or after-treatmentdevice. The primary fuel is diesel, and it teaches using a second fuelsuch as distilled diesel, gasoline, natural gas, liquid petroleum gas(LPG), or hydrogen, which is temporarily injected into the intakemanifold to premix with air before it is introduced into the combustionchamber. In U.S. Pat. No. 7,389,752, SWRI also teach an engineembodiment where gasoline is the preferred primary fuel and lubricatingoil is the micro pilot ignition fuel. A high level of EGR (for instance25-60%) can be used. Neither of these disclosures by SWRI teaches usinga gaseous fuel as the primary fuel, and adjusting the method ofoperating the engine in a different way from a conventionalliquid-fuelled engine to take advantage of the different properties ofgaseous fuels such as, for example, the combustion of such gaseous fuelsproducing less particulate matter, also known as soot, which can allowhigher levels of EGR without the effect of recirculating large amountsof soot, and the generally higher flammability limits and longerignition delays that can help to reduce the danger of combustionknocking.

Even though internal combustion engines have undergone continuousimprovement for more than a century, the combustion process in aninternal combustion engine is complex and even now it is not fullyunderstood. There are many variables and combinations of features thathave not been tried and without investigation by computer modelingand/or experimental testing, the effect of a previously untriedcombination can not be accurately predicted. As discussed above, withrespect to gaseous-fuelled engines, there have been approaches that haveused spark ignition, pilot ignition, hot surface ignition, and therehave been lean burn engines, stoichiometric engines, and there have beenport injected fumigated engines with pre-mixed fuel-air mixtures anddirectly injected stratified fuel-air mixtures, and there have beenengines that use three way catalysts and engines that use relativelymore complex aftertreatment subsystems such as selective catalyticreduction, which requires the addition of a reductant such as urea.

A concern with engine technology in general is the need to preventunacceptable combustion knock which can become more problematic asin-cylinder temperatures get higher and/or with higher compressionratios and/or lower octane fuels, and so on. Various techniques havebeen suggested in the art to control or reduce combustion knock. Forinstance, U.S. Pat. No. 7,028,644 discloses adding hydrogen to avoidcombustion knocking and to allow for higher levels of cooled EGR inspark ignited, gasoline engines with high compression ratios. U.S. Pat.Nos. 7,290,522 and 7,461,628 disclose two mode engines with addition ofhydrogen or varied amounts of injected ethanol to respectively preventcombustion knock.

Much work has been done to improve engine performance and provide foralternative fuel use. Among known gaseous-fuelled engine technologies,HPDI has been shown to yield the highest performance and efficiencies,which makes HPDI the preferred choice for certain applications. However,for less demanding applications, which do not require such highperformance, there is a need for an engine that is simpler and lessexpensive. The present technique addresses this and other needs.

SUMMARY OF THE INVENTION

The present technique relates to gaseous-fuelled stoichiometriccompression ignition, internal combustion engines that are operated withpilot ignition. According to the present method, the primary fuel is agaseous fuel with the pilot fuel being a fuel that is more readilyauto-ignited under the normal conditions found in a compression ignitionengine. The total amount of fuel is the amount of gaseous fuel combinedwith the amount of pilot fuel. On average, on an energy basis, thegaseous fuel, being the primary fuel, represents the majority of thefuel consumed by the engine, and depending upon the operating conditionsthe primary fuel can be at least up to 90% of the total fuel deliveredto the combustion chamber. The method generally comprises determining bymass the amount of primary fuel introduced into the combustion chamberbased on a desired engine load, taking into account the energy providedby the pilot fuel so as to not result in over-fuelling. Also based uponthe desired engine load and/or other engine operating conditions, themethod further comprises determining by mass an amount of exhaust gasthat is cooled and recirculated back to the combustion chamber throughan exhaust gas recirculation subsystem. Then, with the total amount offuel known, and the amount of exhaust gas to be recirculated also known,an amount of air from an intake manifold is delivered into thecombustion chamber, wherein the amount of air is controlled using anair/fuel ratio control means and the predetermined amount of air ismatched to the total amount of fuel to produce essentiallystoichiometric conditions within the combustion chamber during normaloperating conditions. The pilot fuel is injected directly into thecombustion chamber through a pilot fuel injector, with the timing forinjecting the pilot fuel being late in the compression cycle, takinginto account the ignition delay for the pilot fuel, and timing the startof combustion to occur at or near top dead center.

The present method uses the determination of the amount of fuel andexhaust gas by mass, and a person skilled in the art will easilyunderstand that equivalent methods of determining the above amount bymass can be based on measurements of volume and pressure or otherparameters that correlate to mass.

While the pilot fuel injector is located where its nozzle can injectpilot fuel directly into the combustion chamber, the primary fuelinjector can be one injector with a nozzle for introducing the gaseousprimary fuel into the intake manifold or alternatively directly into thecombustion chamber at low pressure, or the apparatus can comprise aplurality of primary fuel injectors, each one associated with arespective intake port or combustion chamber. When the gaseous fuel isinjected directly into the combustion chamber, the timing of the gaseousfuel injection is preferably controlled to take place during thebeginning of the compression stroke, for example the gaseous fuelinjection is controlled to start between 80 and 180 crank angle degreesbefore top dead centre. Also, the gaseous fuel injection can becontrolled to start during the intake stroke.

In the present method, the air/fuel ratio control means can comprise athrottle, different variable valve actuation strategies, pulse widthmodulation of the natural gas injected into the intake manifold,controlling the EGR valve in the exhaust gas recirculation system, or acombination thereof. By “pulse width” a person skilled in the artunderstands the duration of a gaseous fuel injection event.

A person skilled in the art would understand that some variable valveactuation (“VVA”) strategies comprise varying only the timing of theintake valve opening and/or closing while other actuation strategies cancomprise only varying the intake valve lift and still other strategiescan allow varying both the timing and the lift of the intake valve.

In preferred embodiments, because the primary fuel is pre-mixed with theintake air, and the engine's compression ratio is kept high enough forreliable compression ignition of the pilot fuel, the present engine andmethods for operating it preferably employ strategies designed toprevent combustion knock. To control combustion knock, the method cancomprise one or more steps selected from the group consisting of:choosing an engine compression ratio that is lower than that ofconventional diesel engines, but higher than that of conventional SESIengines, dynamically controlling the compression ratio by varying timingfor opening and closing the intake valve for each cylinder, increasingthe cooling of the intake air such that the intake charge mixture isbelow 60° C. when it enters the combustion chamber; increasing theamount of recirculated cooled exhaust gas; increasing the cooling of therecirculated exhaust gas; and managing ring blow-by and positive crankcase ventilation so as to be less than about 2% of the total flow to theengine. The engine preferably operates at a compression ratio smallerthan 14:1.

The present engine and methods for operating it can also employstrategies to control the pumping losses when the engine is operatingbelow a predetermined low load by controlling the timing of the intakevalve opening and closing and/or the intake valve lift, deactivating atleast one of the engine's combustion chambers such that air and gaseousand pilot fuels are introduced in a reduced number of engine combustionchambers, or opening a bypass valve in the engine's exhaust gas systemsuch that at least some of the exhaust gas bypasses a turbochargerinstalled on an exhaust line connected to said engine.

One preferred VVA strategy which involves early intake valve closing(“EIVC”), comprises opening the intake valve when the piston associatedwith the engine combustion chamber is at or near top dead center (“TDC”)of the intake stroke and closing the intake valve when before the pistonreaches bottom dead centre (“BDC”) of the intake stroke and preferablybefore the piston reaches 20 crank angle degrees before the bottom deadcentre of the intake stroke.

Another preferred VVA strategy which involves late intake valve closing(“LIVC”) comprises opening the intake valve when the piston associatedwith the engine combustion chamber is at or near top dead centre of theintake stroke and closing the intake valve during the piston'scompression stroke when the piston is at a crank angle greater than 20degrees after bottom dead centre of the piston's intake stroke.Preferably, the intake valve is closed before the piston reaches 120crank angle degrees after the bottom dead centre of the intake stroke.

The present engine generally comprises a combustion chamber; a gaseousprimary fuel injector for introducing gaseous primary fuel into anintake air manifold, or an intake port or directly into the combustionchamber; a pilot fuel injector for injecting pilot fuel into thecombustion chamber; an exhaust gas recirculation subsystem; means forcontrolling the air/fuel ratio of the combustion mixture; and an enginecontroller programmed to control the gaseous primary fuel injector, thepilot fuel injector, the exhaust gas recirculation subsystem, and theair/fuel ratio control means such that the air/fuel ratio is essentiallystoichiometric during normal operation of the engine.

The gaseous primary fuel injector is preferably located in the intakeair manifold or an intake port, and for engines having a plurality ofcylinders, it can be desirable to employ a plurality of primary fuelinjectors, with one associated with each intake port. This approach isknown as port injection. In another embodiment, the gaseous primary fuelcan be introduced directly into each cylinder with the introduction offuel timed to be during the intake stroke or early in the compressionstroke so that it is still injected at relatively low pressure, comparedto more conventional direct injection approaches that introduce the fuellater in the compression stroke when the piston is closer to top deadcenter.

The means for controlling the introduction of air and/or recirculatedexhaust gas into the combustion chamber comprise throttling means, anintake valve, an exhaust valve and an EGR valve. The engine comprises acontroller programmed to control the means for controlling theintroduction of air and/or recirculated exhaust gas into said combustionchamber to achieve a stoichiometric oxygen/fuel ratio during theengine's normal operation. The controller is programmed to control theintake valve opening and closing and/or the intake valve lift to achievea stoichiometric oxygen/fuel ratio.

In one preferred embodiment, the engine controller is programmed tocontrol the intake valve such that it opens when a piston associatedwith the combustion chamber is at or near top dead centre of thepiston's intake stroke and it closes during the piston's intake stroke.Preferably the controller controls the intake valve such that it closesbefore 20 crank angle degrees before bottom dead centre of the piston'sintake stroke.

In another preferred embodiment, the engine controller is programmed tocontrol the intake valve such that it opens when a piston associatedwith the combustion chamber is at or near top dead centre of thepiston's intake stroke and it closes during the piston's compressionstroke, at a crank angle greater than 20 degrees after bottom deadcentre of the piston's intake stroke. Preferably the controller isprogrammed to control the intake valve such that it closes before 120crank angle degrees after bottom dead centre of the piston's intakestroke.

In one embodiment of the present engine an exhaust gas line connected tothe engine comprises a turbocharger and a bypass valve for bypassing theturbocharger and the engine controller controls the bypass valve tocompletely or partially open it when the engine operates at low loads.This has the effect of lowering the pumping work required for theengine's efficient operation.

The engine controller is also programmed to deactivate at least one ofthe engine's cylinders such that air and gaseous and pilot fuels areintroduced into a reduced number of cylinders for generating powerthrough combustion when the engine is operating at low loads. This alsoreduces the overall pumping work required for the engine's efficientoperation.

An advantage of the present arrangement is that the internal combustionengine can comprise an inexpensive three way catalyst exhaust treatmentsubsystem. The means for controlling the air/fuel ratio of thecombustion mixture can comprise a throttle, different variable valveactuation strategies or pulse width modulation (PWM) of the gaseous fuelinjector.

The present stoichiometric, compression ignition, internal combustionengine and this operation strategy has been found to be particularlysuited for operation with natural gas as the gaseous primary fuel anddiesel as the pilot fuel.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic view of a throttled, stoichiometric, compressionignition, internal combustion engine equipped with a three way catalystexhaust treatment subsystem and having an injector that injects gaseousfuel into the intake manifold.

FIG. 2 shows the intake valve lift and opening duration for differentvalve actuation strategies that employ early or late intake valveclosing.

FIG. 3 is a schematic view of throttled, stoichiometric, compressionignition, internal combustion engine equipped with a three way catalystexhaust treatment subsystem and having an injector that injects gaseousfuel directly into the combustion chamber.

FIGS. 4 a, b, and c show indicated mean effective pressure (IMEP), knockindex, and thermal efficiency versus start of combustion timing resultsrespectively for exemplary SESI and the present engines in thecalculated engine examples.

FIG. 5 shows the modeled specific fuel consumption (ISFC) versus thebrake mean specific pressure (BMEP) for different intake valve closingtimings applied to a medium duty gaseous fuelled internal combustionengine.

FIG. 6 shows the modeled in-cylinder temperature during the compressionstroke when employing different intake valve closing timings for amedium duty gaseous fuelled internal combustion engine.

FIG. 7 shows the modeled net specific fuel consumption for differentengine operation modes employing throttling, turbocharger bypass and/orvariable valve actuation strategies.

DETAILED DESCRIPTION OF PREFERRED EMBODIMENT(S)

For purposes of focus and brevity certain aspects of the embodimentsthat are conventional and well known to those familiar with gaseousfuelled engine technology are not described in detail.

In the prior art, certain terms have been used in different contexts andin different publications to have different meanings. Accordingly, inthe context of this disclosure and the description of the present methodand apparatus, the following terms are defined as follows:

“Stoichiometric” refers to situations pertaining to oxidant and fuelmixtures in which the total amount of oxidant is balanced with the totalamount of fuel present such that both would be essentially completelyconsumed when combusted. Numerically the total oxidant/total fuel ratiois preferably 1±0.1 and more preferably 1±0.05.

“Normal” engine operation refers to the various routine modes ofoperation for the engine when operating in a steady state (for example:idle, light load, full load). Most of the time, the engine is expectedto operate in one of these modes. However, this does not include specialsituations, such as relating to maintenance, diagnostics, transientconditions or the like.

“Low pressure” with respect to fuel injection refers to pressures belowabout 50 bar.

FIG. 1 shows a schematic view of engine system 1 which comprises anexample of one embodiment of a gaseous fuelled stoichiometric,compression ignition, internal combustion engine. In this example,stoichiometry can be controlled in part using throttle 20 to control theamount of intake air and in part by using EGR valve 15 to control theamount of exhaust gas that is recirculated. Engine system 1 is alsoequipped with three way catalyst 21 for treating the engine exhaust thatexits combustion chamber 4 through exhaust valve 6.

An actual engine typically comprises a plurality of cylinders andcombustion chambers, but for illustrative purposes the internalcombustion engine shown in FIG. 1 only comprises combustion chamber 4defined by cylinder 2, and reciprocating piston 3. Intake valve 5 andexhaust valve 6 respectively allow for the entry of the air/fuel mixtureinto combustion chamber 4 and for the exit of combustion exhaust gasestherefrom. Engine system 1 is also provided with intake manifold 7 andexhaust manifold 8 that are fluidly connected to intake valve 5 andexhaust valve 6 respectively.

The primary fuel used in engine system 1 is a gaseous fuel, which issupplied from primary fuel supply 9 and introduced into intake manifold7 by primary fuel injector 11 in the vicinity of intake valve 5. Theshown embodiment is sometimes described as port injection because thatportion of intake manifold 7 is typically described as the intake port.Relatively low injection pressures, for example about 1-15 bar (gauge),can be used when introducing the gaseous fuel at this location. Sinceport injection is well known for gaseous fuel such as natural gas,primary fuel supply 9 can be an appropriate fuel supply subsystemdeveloped for supplying port injectors with the fuel stored as eithercompressed or liquefied natural gas.

Compression alone is generally not a reliable approach for ignitingair/gaseous fuel mixtures. As previously discussed with respect to knownapproaches, a strategy for assisting the ignition of an air/gaseous fuelmixture is desirable. Whereas with port injected gaseous fuels thisignition assist is normally achieved with a spark plug, the presentarrangement employs an appropriate pilot fuel, such as diesel fuel,which is supplied from pilot fuel supply 10 via a small common rail andintroduced directly into combustion chamber 4 by pilot fuel injector 12.Relatively high injection pressures, for instance 200-2000 bar, are usedto atomize and disperse the liquid diesel fuel and to overcome therelatively higher pressures present in combustion chamber 4 as a resultof the compression stroke, compared to the pressures in intake manifold7.

Engine system 1 further comprises an exhaust gas recirculation (EGR)subsystem represented schematically here by EGR conduit 14 andcontrollable EGR valve 15. As shown in FIG. 1, the amount of ambient airadmitted to intake manifold 7 is regulated by throttle 20. Electronicprogrammable engine controller 16 receives signals from sensors (notshown) that detect certain engine operating parameters such astemperature, pressure, fuel levels in the fuel storage tanks, and thatindividually or collectively indicate certain engine conditions andcontrols various engine components, some of which are shown, such asprimary fuel injector 11, pilot fuel injector 12, EGR valve 15, intakevalve 5, exhaust valve 6 and throttle 20. In FIG. 1 the dashed linesrepresent the signal wires that send sensor signals to controller 16 andthe signal wires that send control signals from controller 16 to theconnected engine components. It is generally recognized that enginecontroller 16 can be used to sense other engine operating parameters andto control components other than those depicted here.

In normal operation, engine controller 16 controls engine components inresponse to a variable engine load determined by an operator. Controller16 determines a desired stoichiometric combustion mixture based on acommanded load and increases or decreases the injected amount of gaseousprimary fuel accordingly. In addition, controller 16 adjusts throttle20, intake valve 5 and EGR valve 15 to admit desired amounts of air anddiluting exhaust gas respectively into intake manifold 7. Later, duringthe compression stage, controller 16 operates pilot fuel injector inorder to obtain the desired amount, rate, and timing of injected pilotfuel.

During normal operation engine controller 16 is programmed to maintain astoichiometric air/fuel ratio and if controller 16 detects that theair/fuel ratio has strayed from being stoichiometric, controller 16controls the connected engine components to restore a stoichiometricoxidant/fuel ratio. As described herein, controller 16 can have aplurality of components that it can control in order to adjust theoxidant/fuel ratio and controller 16 is programmed to adjust one or moreof such components depending upon the detected engine parameters and thedetermined engine conditions. By maintaining a stoichiometric air/fuelratio, the engine exhaust is substantially oxygen free and thus arelatively simple and inexpensive conventional three way catalystsubsystem 21 is employed as an emissions treatment subsystem. However,without departing from the present technique, for brief periods, engine1 can be operated in non-stoichiometric modes such as during cold start,regeneration of exhaust gas treatment devices, when operating in anexhaust gas catalyst temperature protection mode, or a diesel only mode(for example, when natural gas is not available.

Instead of or in addition to employing throttle 20, an alternativeembodiment for controlling the air/fuel mixture uses different variablevalve actuation strategies to control the amount of charge introducedinto the combustion chamber through intake valve 5.

In the shown embodiment, the opening and closing of intake valve 5 andexhaust valve 6 is controlled by controller 16. In preferredembodiments, variable valve actuation strategies enable control of bothtiming and lift of the intake valve. Such variable valve actuationdevices comprise electronically controlled electromagnetic valves.Variable valve actuation strategies have introduced an additionalvariable to engine operation that has unexpected effects. For example,from experimental results it was found that the strategies that employjust advancing or delaying the timing of the intake valve openingcompared to a conventional intake valve opening timing, withoutcontrolling how long the intake valve stays open, did not have abeneficial effect. Such timing strategies either increase the pumpingwork required to deliver the necessary amount of air to the combustionchamber or result in a net increase in fuel consumption, as it wasobserved when delaying the timing of the intake valve opening, or theycan interfere with the piston stroke and can introduce relatively highlevels of residuals into the exhaust stream, as it was observed whenadvancing the timing of the intake valve opening. It was found that itis more beneficial for the overall engine efficiency to maintain thetiming of the intake valve opening at around TDC (top dead centre)before the piston's intake stroke and to advance or delay the timing ofthe intake valve closing, as further described below.

FIG. 2 illustrates the intake valve lift and the intake valve openingduration, measured in crank angle degrees “after top dead centre firing”(ATDC_(f)) for different intake valve actuation strategies. The crankangle is zero at TDC_(f) which indicates the point at which the cylindervolume is minimized between the piston's compression and power strokes.In FIG. 2 a crank angle of 360° ATDC_(f) indicates top dead centre (TDC)which is the point of minimum cylinder volume at the beginning of thepiston's intake stroke and a crank angle of 540° ATDC_(f) indicatesbottom dead centre (BDC) which is the point of maximum cylinder volumeat the end of the piston's intake stroke.

A conventional engine which does not have variable valve actuationtypically has a valve lift profile as the one indicated by referencenumber 100 with the intake valve opening at TDC and closing at or nearBDC. A first VVA embodiment employing a variable valve actuationstrategy is indicated by the intake valve lift profile 200 and comprisesopening the intake valve around the same time as with a conventionalstrategy, more specifically around top dead centre (TDC) before theintake stroke, at a crank angle of 360° ATDC_(f) and closing the intakevalve early compared to the conventional strategy indicated by theintake valve lift profile 100 for the same engine. That is with thisfirst embodiment the intake valve is closed during the piston's intakestroke before the piston has moved to a BDC position. This strategypreferably involves reducing the intake valve lift.

The early intake valve closing illustrated by the intake valve liftprofile 200 comprises closing the intake valve at least 20 crank angledegrees before the intake stroke ends (before BDC). Preferably, theearly intake valve closing illustrated by profile 200 comprises keepingthe intake valve open for 20 to 180 crank angle degrees after openingit.

A second VVA embodiment indicated by the intake valve lift profile 300comprises opening the intake valve around the same time as with aconventional strategy indicated by the intake valve lift profile 100,more specifically at around TDC before the intake stroke, at a crankangle of 360° ATDIC_(f), and closing the intake valve later during thecompression stroke, more specifically closing the intake valve after 20crank angle degrees after the end of the intake stroke (after BDC), at acrank angle greater than 560° ATDIC_(f) Late intake valve closingcomprises closing the intake valve at around 20 to 120 crank angledegrees after the intake stroke ends (after BDC). In this second VVAembodiment, the maximum intake valve lift could be less than the maximumintake valve lift used in the conventional strategy indicated by theintake valve lift profile 100 for a portion or all of the duration ofthe intake valve opening.

The inventors' experimental results have shown that better fuelconsumption and emission reduction at both low and high loads can beachieved by the early and late intake valve closing strategies describedabove.

During engine operation at high loads both early intake valve closingand late intake valve closing have the effect of reducing the effectiveengine compression ratio with an overall result of reducing thein-cylinder temperature at the end of compression. By closing the intakevalve early or late it was observed that the danger of engine knock athigh loads can be reduced without having to alter the engine's geometriccompression ratio. That is, the present method allows engine operationat higher geometric compression ratios than conventional SESI engines,while operating with lower effective compression ratios which has anoverall result of reducing the in-cylinder temperature at the end ofcompression and preventing knock. This is especially beneficial forengines fuelled with a stoichiometric fuel ratio, which are more proneto engine knock compared to engines that are fuelled with a lean fuelmixture. Because the compression process inside the combustion chamberessentially starts when the intake process ends, the effectivecompression ratio is defined as the ratio between the volume of thecombustion chamber when the intake air flow through the intake valvestops and the volume of the combustion chamber at the end of thecompression stroke. For both the early and the late intake valve closingthe volume of the combustion chamber at the time when the intake airflow through the intake valve stops is smaller than the volume of thecombustion chamber when the piston reaches the bottom dead centre(before the compression stroke), as is the case in the conventionalintake valve lift profiles, and therefore the effective compressionratios for the early and late intake valve closing are smaller than theconventional effective compression ratios. This leads to lowerin-cylinder temperatures at the end of the compression stroke and in theend gases during combustion, reducing the potential for knock.

To maintain the required power for the engine's stoichiometric operationthe charge mass introduced into the combustion chamber during the abovedescribed valve actuation strategies has to be maintained relativelyconstant to maintain a given power output. This can be achieved bycontrolling the boost pressure which is defined to be the pressure inthe intake manifold. As the effective compression ratio is reduced, theboost pressure required to meet a given load increases.

While at high loads the above valve actuation strategies do notgenerally require the use of throttling, at low loads, throttling ispreferably used together with varying the timing of the intake valveclosing to adjust the amount of air introduced in the combustionchamber, especially during short intake valve opening durations. Theintake valve response time is generally slower than the throttleresponse time and therefore it would be generally more difficult tocontrol the air mass introduced into the combustion chamber just byusing only the above described valve actuation techniques at low loads.

Another technique for preserving the required charge mass to be injectedinto the combustion chamber at low loads while reducing pumping lossesis deactivating some of the engine cylinders so that the engine operateswith a reduced number of cylinders. Cylinder deactivation involvesincreasing the charge mass that is introduced into the combustionchamber per firing cylinder; this results in a higher in-cylinderpressure during the intake stroke for the firing cylinders and thereforereduced pumping work.

Engine systems generally comprise a turbocharger, where a turbineinstalled in the exhaust system uses the enthalpy in the exhaust todrive a compressor in the intake system which increases the intake airpressure. For such systems, another technique for maintaining therequired charge mass introduced into the combustion chamber whilereducing the pumping work done by the system at low loads is to open abypass valve (called “wastegate”) such that some or all of the exhaustgas bypasses the turbocharger after exiting the engine. When bypassingthe turbocharger less backpressure is created in the exhaust gas lineand therefore less pumping work is required for pushing the exhaust gasout of the combustion chamber. Also, when some or all of the exhaust gasflow bypasses the turbocharger, the pressure of the air supplied to theengine remains lower and intake air can be delivered to the enginewithout throttling.

In another alternative embodiment, shown in FIG. 3, the low pressureprimary fuel injector 11A can be located with a nozzle located withincombustion chamber 4 with the primary fuel injector valve body mountedin the cylinder or in the cylinder head for introducing the gaseous fueldirected into the combustion chamber instead of with the intake airthrough intake valve 5. In order to avoid the added complexityassociated with raising the gaseous fuel to a higher injection pressure,according to the present method, injection of the gaseous fuel ispreferably completed early during the compression stroke. This approachcan be more advantageous in embodiments that employ variable valveactuation strategies and devices, because then the amount of oxidantintroduced into combustion chamber 4 is controlled more independentlyfrom the amount of fuel. The embodiment illustrated in FIG. 3 has manycomponents that are the same with the components of the embodimentpresented in FIG. 1 and they have been identified with the samereference numbers. These components have been already described withrespect to the embodiment illustrated in FIG. 3 and therefore theirpurpose and function is not repeated here.

Similar variable intake valve actuation strategies to the ones describedabove can be applied to the intake valves used for engines where thegaseous fuel is injected directly into the combustion chamber, and insuch embodiments the closing of the intake valve is preferably timed toprecede the gaseous fuel injection. That is an intake valve profile likethe one indicated by reference 200 in FIG. 2 would be an example of apreferred embodiment since this would facilitate early injection of thegaseous fuel.

A still further alternative embodiment could employ pulse widthmodulation of the natural gas injected into the intake manifold. Pulsewidth modulation is a common, inexpensive and robust way to reliablydeliver fuel to the intake manifold in a precise and controlled way.

The present method and apparatus offers a unique combination ofelements. Whereas each individual element on its own might have expectedbenefits such benefits are not certain or may be limited when combinedwith other elements because a change in one element can affect anotherelement. The present method and apparatus has been found to produceperformance characteristics different from known gaseous-fuelledengines. For example, compared with SESI gaseous-fuelled engines,compression ignition of diesel pilot fuel enables operation at highercompression ratios and increased level of EGR. In turn, this leads toreduced fuel consumption (up to ˜5%), higher BMEP ratings, and higherefficiency than is currently possible in SESI engines (mainly due tocombustion knock limitations). Also, relative to SESI engines, animproved robustness to variations in natural gas quality is expected(since the ignition energy from combustion of the pilot fuel is severalorders of magnitude higher than that of a spark and there are multiplesites for ignition thereby promoting faster and more uniform burning,which in turn is believed to lead to increased EGR tolerance and reducedknock tendency). It is believed that compared to SESI engines, thedurability of the pilot ignition components should be increased becausethe lifetime of injectors is typically several times that of sparkplugs. With regard to emissions, properties can be similar to SESIengines because of the stoichiometric air/fuel ratio maintained duringnormal operation by the present method and apparatus allowing the use ofa conventional 3 way catalyst emissions treatment subsystem, with thepossible exception of particulate emissions due to the use of smallamounts of pilot fuel.

Unlike dual fuel natural gas engines that operate in lean mode (that is,with a significant excess of oxygen in the fuel mixture), the presentmethod and apparatus does not require a more complicated and expensiveemissions treatment subsystem to meet current emissions standards. Atypical dual fuel emissions treatment subsystem employs an oxidationcatalyst (oxi-catalyst), selective catalytic reduction (SCR), and adiesel particulate filter (DPF). Thus, the unique combination ofelements taught by the present method and apparatus offers benefits overdual fuel engines in this regard.

Of prior art engine options, HPDI gaseous-fuelled engines offer superiorperformance, for example, as measured by efficiency and BMEP. However,these engines require a high pressure gaseous fuel delivery subsystem toenable injection of the gaseous fuel directly into the combustionchamber late in the compression stroke, and this generally requires apressure boosting pump and gaseous fuel injectors with nozzles that arelocated inside the combustion chamber. Finding room within thecombustion chamber to locate nozzles for a gaseous fuel injector and apilot fuel injector can be a challenge in small engines, especiallymodern engines with much of the space in the cylinder head alreadyoccupied by two intake valves and two exhaust valves. Furthermore, likeconventional diesel engines, HPDI gaseous-fuelled engines operate with astratified combustion process, so HPDI gaseous-fuelled engines typicallyrequire a more complicated and expensive emissions treatment subsystem,which typically includes an oxi-catalyst, combined with SCR and a DPF.Here again, the different combination of elements taught by the presentmethod and apparatus produces a different result that provides a simplerand potentially less expensive alternative that could be easier toimplement in smaller engines, while still offering competitiveperformance.

Further to what has already been described, Table 1 illustrates some ofthe key differences between the presently present method and apparatuscompared to the above-described known engines and methods of operatingthem, by comparing some of their attributes and/or characteristics.

TABLE 1 Present Method Attribute/characteristic SESI Dual fuel HPDI andApparatus Ignition Spark Compression Compression Compression methodignition of ignition of pilot ignition of pilot ignition of pilotgaseous fuel triggers ignition triggers ignition triggers ignition (nopilot) of gaseous fuel of gaseous fuel of gaseous fuel Ignition Sparkplug Pilot fuel Pilot fuel Pilot fuel hardware injector for injector forinjector for injecting pilot injecting pilot injecting pilot fueldirectly fuel directly fuel directly into the into the into thecombustion combustion combustion chamber chamber chamber Gaseous fuelLow Low pressure, High pressure, Low pressure, injection pressure,single or multi direct in single or multi method single or point inintake cylinder point in intake multi point manifold or manifold or inintake ports ports, or directly manifold or into cylinder ports Expected% 100% Up to ~60% Up to ~90% Up to ~90% gaseous fuel of total fuel (onan energy basis) Throttle or Yes No No Yes other air/fuel ratio controlmeans Main Stoichio- Lean Stratified Stoichiometric combustion metrichomogenous homogeneous homogen- eous Possible ≦12:1 ≦17:1 ≦20:1 ≦14:1compression ratio Relative Moderate Moderate High Moderate to potentialhigh efficiency Potential 20 bar 16 bar (if <10% 30 bar 23 bar BMEP ofenergy comes from pilot fuel Potential for Yes Yes No Yes Engine KnockEmission 3 way Oxi-catalyst, Oxi-catalyst, 3 way catalyst treatmentcatalyst DPF, & SCR DPF, & SCR & DPF Relative Cost Lowest Moderate HighLow

While the unique approach of the present engine arrangement offerssimplifications and/or other advantages over known gaseous-fuelledengines, there are also some aspects that present different challenges.For example, in an HPDI engine, combustion knock is not a problembecause the fuel is generally injected directly into the combustionchamber at a timing that is near top dead center so that there issubstantially no possibility of combustion knock. With the presentengine arrangement, the primary gaseous fuel is introduced into thecombustion chamber with the intake air or earlier in the engine cycle sothat premature detonation, which is what is known as combustion knock isa possibility that needs to be guarded against, especially when it isdesirable to retain a relatively high compression ratio to maintainrobust ignition under all relevant operating conditions.

Reducing the engine's compression ratio, compared to diesel engines, isone approach to reducing the likelihood of combustion knock that istypically used by SESI engines and to a lesser degree by Dual Fuelengines. However, reducing the compression ratio makes it morechallenging to promote strong and reliable ignition under all relevantoperating conditions.

It is estimated that the energy from the ignition of a small amount ofpilot diesel fuel can be three to four orders of magnitude greater thanthat of a spark plug in a SESI engine, and by using a plurality of pilotfuel injection sprays, a much larger ignition zone is provided by thepilot fuel. This results in improved ability to burn a charge that isdiluted with higher levels of EGR, so that unlike SESI engines, whichare limited by the ignition energy of the spark plugs, with the presentmethod and apparatus more EGR can be used to suppress engine knock. Atthe same time, higher levels of EGR also affect the air/fuel ratiobecause there is substantially no oxygen in the recirculated exhaust gaswhich displaces some of the fresh air, so the amount of EGR commandedneeds to be coordinated with the control of throttle 20.

In the examples presented below, combustion computational fluid dynamicsmodeling was carried out to investigate knock tendency, efficiency andemissions of the present method and apparatus. These simulations showthat the present method and apparatus can achieve a remarkably high BMEP(>23 bar) and can moderately improve thermal efficiency compared toconventional SESI engines, while still acceptably controlling combustionknock. Combustion in the present engine yields a higher rate of heatrelease than a conventional SESI engine so that the optimal ignitiontiming for best thermal efficiency is retarded compared to a SESIengine. This retarded start of ignition timing allows the presentengines to operate at higher brake mean effective pressure (BMEP) withsignificantly lower combustion knock indices than otherwise identicalSESI engines. This advantage is more pronounced at higher compressionratios.

To further reduce the tendency for combustion knock, the present methodand apparatus also allows the use of several additional strategies.Because combustion knock is most sensitive to the in-cylindertemperature in the engine, a strategy that can reduce the end gastemperature can be beneficial. These strategies include increasing theamount of cooled EGR to increase charge dilution, sizing the EGR coolerto provide increased cooling to the EGR gases, increasing intake chargeair cooling such that intake charge temperature on entering the cylinderunder all relevant operating conditions is below 60° C., lowering theengine effective compression ratio by varying valve actuation asdescribed earlier. A further strategy includes reducing ring blow-by tobe less than about 2% relative to the airflow on a volumetric basis.(This is because oil in the blow-by contributes to combustion knockreactions.)

In order to suppress combustion knock with the present method andapparatus, computational and experimental results suggest that it isbest to keep the compression ratio lower than that of a conventionaldiesel or HPDI engine. As a result, it is more difficult to ensure thatthe injected pilot diesel fuel will auto-ignite under all relevantoperating conditions. Calculations show that the present enginearrangement can theoretically ignite diesel at compression ratiosbetween 12:1 and 13:1 for engine operating conditions above 0° C. Thesecalculations involved determining the effect of compression ratio on endof compression temperature in the engine cylinder for various intakecharge temperatures. Below 0° C., a cold starting device such as a blockheater or an air heater may be needed; using a glow plug is notrecommended in this case as it has the potential to induce the charge toignite during the intake event, leading to uncontrolled combustion inthe intake manifold (“backfire”). Methods of controlling the actuationof the intake or exhaust valve could also be used to increase thein-cylinder temperature and thereby enhance auto-ignition of the dieselpilot, by retaining more hot exhaust gases in the cylinder.

Other technical issues may require special consideration when usingembodiments of the present engine. For example, as previously discussed,while the present engines can be more tolerant to variations in naturalgas composition compared to SESI engines, the present engines can beless tolerant in this respect compared to HPDI gaseous-fuelled engines.Accordingly, in some embodiments combustion sensing and a controlstrategy for making adjustments to engine operation responsive tomeasured combustion behavior can be helpful in combination with thepresent engine arrangement. For example, co-owned U.S. Pat. Nos.7,133,761, 7,200,487 and 7,444,231 disclose examples of methods andapparatuses that could be used for this purpose. Also, if micro-pilotfuel operation is employed, attention will be required in order tocontrol the pilot fuel quantity to sufficiently small levels whileretaining adequate pulse width and sufficiently high injection pressure,for example, to aid atomization. Furthermore, due to the considerablyreduced flow of pilot fuel in some circumstances, the pilot fuelinjector is preferably designed to operate satisfactorily at highertemperatures than conventional diesel injectors and/or be designed withfeatures for cooling the injector tip or other features for preventingoverheating and/or the accumulation of carbon deposits, also known as“tip carboning”.

Another technical challenge relates to accurate control of the enginesystem, which is expected to be more complicated than that for a SESIengine due to the presence of pilot diesel. Nevertheless, while uniquecontrol strategies are needed for the present method and apparatus,technology already exists for managing two different fuels, such as thetechnology developed for HPDI gaseous fuelled engines that use a primarygaseous fuel and a pilot fuel for ignition. And likewise, whileembodiments of the present method and apparatus can incorporate similarsteps and components as those for prior art engines, to reflect theuniqueness of the present combination of elements the combination ofsteps and the manner of controlling the different elements aredifferent. For example standard methods can be used in such things asthe determination of the optimal compression ratio (to achieve pilotignition under typical operating conditions while suppressing combustionknock and achieving highest possible BMEP and lowest possible methanenumber operation), preferred injector design (for instance number, shapeand configuration of jet holes), and piston bowl shape and swirl ratio(note that the reverse nebula piston shape used in some SESI engineswill most likely not be suitable for the present engine configurationdue to interference of pilot fuel spray by the piston shape).

The following examples are provided to illustrate certain aspects of thepresent method and apparatus, but should not be construed as limiting inany way.

Calculated Engine Examples:

A numerical study was carried out to investigate combustion knocktendency and efficiency of engines of the present technique and of SESIengines operating under comparable conditions. The study involved use ofa two-zone premixed combustion model and was focused at a high BMEPrange (20-30 bar). The two-zone combustion model is described inCatania, A. E., Misul, D., Mittica, A. and Spessa, E., “A RefinedTwo-zone Heat Release Model for Combustion Analysis in SI Engines”, JSMEInternational Journal, Vol. 46, No. 1, 2003, and was modified to explorea wider parameter space to evaluate both the present combustion strategyand spark ignition combustion. The propagation of flame was modeledbased on the laminar flame velocity calculated from local mixturetemperature, pressure, and turbulence wrinkle factor determined from anempirical model. Diesel injection and combustion was modeled byaccelerating the propagation of flame within the projected volumecovered by the diesel spray. At given conditions, the potential knockintensity was characterized by the projected rate of pressure rise inthe end gas. The knock intensity estimated in this example representspotential rate of energy release from the end gas assuming that the endgas is consumed within one characteristic time scale. The characteristictime scale is the auto-ignition delay time of the end gas.

The engine geometries, operating conditions and parameters used in thisstudy are summarized in Table 2 below.

TABLE 2 Engine and Operating Parameters Two-Zone Model EngineDisplacement (L) 9 Methane Numbers 87, 50 Swirl Ratio 1.0 PilotQuantities (as % of total fuel 2-2.5 energy) Engine Compression Ratios11:1, 14:1, 18:1 Diesel Injection Timings (degrees 30, 22, 15, 10, 5 CABTDC) Intake Temperatures (° K.) 320, 350 EGR (% of the total mass 21comprising the intake air and EGR) Engine Speed (RPM) 1200 Engine Load(IMEP, bar) 20

From this detailed study, the following exemplary results were obtained.Table 3 below shows a comparison of indicated mean effective pressure(IMEP) in bar for both engine types fuelled with natural gas fuel havinga methane number of 87, at about peak thermal efficiency (that is,T_(intake)=350° K) with the piston design used in the Cummins WestportISL-G commercial engine and ignition timings of 22° and 5° crank anglebefore top dead center (CA BTDC) for the SESI and the present enginesrespectively.

TABLE 3 IMEP (bar) of present Compression ratio SESI IMEP (bar)technique 11:1 19.4 19.5 14:1 19.9 21.1 18:1 20.0 20.6

In addition, FIGS. 4 a, b, and c show IMEP, knock index, and thermalefficiency versus start of combustion timing results respectively forboth a SESI and an embodiment of the present engine operating undersimilar conditions, namely an intake temperature of 320° K, compressionratio of 14:1 and methane number of 87.

As is apparent from FIGS. 4 a, 4 b and 4 c, the present engine exhibitsan improved IMEP and thermal efficiency over a wide range of ignitiontimings. Peak efficiencies for the present engine and a SESI engine inthis example occur at start of combustion timings of 10° and 22° CABTDC, respectively. At these timings the knock index of the SESI engineis better than that of the present engine. However, at 5° CA BTDC, theIMEP and thermal efficiency of the present engine are still close to itspeak values and still better than the peak SESI engine values and theknock index for the present engine is markedly lower than that of theSESI engine at its peak. Furthermore, while more conditions were studiedthan those shown in FIGS. 4 a through 4 c, in the studied conditions,the results obtained were qualitatively the same. That is, the presentengine showed peak IMEPs at 10° CA BTDC but had worse knock index thanthe SESI engine at its peak IMEPs. However at 5° CA BTDC, in every case,the present engines had an IMEP that was close to its peak value, andstill better than the comparable SESI engine peak, while the knock indexwas substantially better than the SESI engine at its peak.

These experimental results show that the present engine and method canunexpectedly provide better IMEP and thermal efficiency in combinationwith reduced knock tendency under the studied operating conditions.

From the studied conditions, it is believed that the present engine isgenerally able to achieve high IMEP and moderately improve on thermalefficiency compared to comparable SESI engines. Combustion in thepresent engine yields a higher rate of heat release than in a SESIengine so the optimal ignition timing for best thermal efficiency isretarded for the former than for the latter. This retarded start ofignition timing allows the present engine to operate at higher IMEP andefficiency with significantly lower knock indices than otherwise similarSESI engines. This advantage is more pronounced at higher compressionratios. With regards to emissions, the study shows that the presentengines potentially have higher levels of NO_(x), carbon monoxide (CO),and particulate matter (PM) exiting the combustion chamber compared toconventional SESI engines operating under similar conditions. But afterconventional 3 way catalyst emissions treatment, based on the knownperformance characteristics of such treatment subsystems, compared to aSESI engine it is believed that the tailpipe emissions levels from thepresent engine system will be substantially the same or lower. Inaddition, with respect to emissions of unburned hydrocarbons, theresults from the computational models show that the difference betweenthe present engine and a SESI engine are insignificant.

Modeling Results

Modeling work done on a 6 cylinder medium duty engine, equipped with avariable geometry turbocharger, an exhaust gas recirculation loop, andprovided with a single injector in the manifold (single point injection)has demonstrated that both early and late intake valve closing bringimprovements to the overall engine efficiency and fuel consumption ascompared to an engine operated with a conventional intake valve liftprofile, such as the one indicated by reference number 100 in FIG. 2.Modeling results illustrated in FIG. 5 demonstrate savings in fuelconsumption over a broad range of BMEP values for both the early intakevalve closing (EIVC) and late intake valve closing (LIVC) compared to amore conventional valve operation with the same engine.

FIG. 6 further illustrates the in-cylinder temperature increase duringthe compression stroke for different intake valve actuation techniques.During the compression stroke, when the volume of the combustion chamberis reduced as the piston moves away from the bottom dead centre (BDC),the in-cylinder temperatures for both EIVC and LIVC are lower at the endof the compression stroke than the in-cylinder temperatures at the endof the compression stroke for the standard engine operation. Thein-cylinder temperature for both EIVC and LIVC at the end of thecompression stroke are equivalent to the in-cylinder temperature at theend of the compression stroke recorded when the engine operates at a lowcompression ratio (for example, 12.7:1 compared to 14:1 which was thecompression ratio used for the standard, EIVC and LIVC operation). Inthis simulation, the EIVC model involved closing the intake valve ataround 500° ATDC_(f) and the LIVC model involved closing the intakevalve at around 630° ATDC_(f).

The modeling results illustrated in FIG. 7 show that operating theengine with variable valve actuation (VVA) in combination with bypassingthe turbocharger achieves the best fuel consumption results, especiallyfor lower BMEPs. Using the turbocharger bypassing or the variable valveactuation techniques separately also achieve an improvement in fuelconsumption over a broad range of BMEP values.

Actual Engine Examples:

A research engine was converted from a 6 cylinder, 15 L engine tooperate on a single cylinder (thus 2.5 L in size). This engine wasoperated in different modes with natural gas as the primary fuel andconventional diesel fuel injected to act as the pilot. The engine wasoperated in stoichiometric mode under the following conditions:

TABLE 4 Speed: 1200 rpm GIMEP (gross indicated mean effective pressure):10.5 bar EGR (% of the total mass comprising the intake air and EGR):30% EQR (fuel/air equivalence ratio): 1 IMT (intake manifoldtemperature): 70° C. GRP (gas rail pressure): 26 MPa Diesel flow: 7 or15 mg/injection (representing 5 and 10% of total energy per injection)Pilot start of injection (PSOI): 20°, 25° or 30° CA BTDC Gas start ofinjection (GSOI): 80°, 100°, 120°, 140°, 160°, or 180° CA BTDC

The research engine was successfully operated under stoichiometricconditions at low to mid-speed and low- to mid-range load conditionswithout hardware changes to the engine.

This is a non-optimized research engine and did not use a natural gasport injector. Further, higher friction is encountered in this 6cylinder engine operating on only 1 cylinder than would be expected in aproduction engine. Accordingly, there are difficulties in estimatingresults expected in a production engine. Still, this engine was able toachieve a thermal efficiency GISFC (gross indicated specific fuelconsumption) between about 200 and 250 g/kW-hr over a 50% IHR(integrated heat release) range from 0 to 12 crank angle degrees aftertop dead center (CA ATDC). Using these values of thermal efficiency andassuming an expected 3% improvement would be obtained in a productionengine, a brake efficiency for a production engine operating in such astoichiometric mode was estimated to be 28%. Accordingly, it is believedthat the experimental results show that competitive brake efficienciescan be obtained when operating in the described stoichiometric modeunder practical conditions.

On the basis of data obtained, the following general observations werealso made:

-   -   combustion timing is mainly determined by diesel injection        timing    -   gas injection timing mainly determines the completeness of the        mixing process    -   increasing diesel flow rate moves combustion timing earlier    -   higher diesel flow rates help improve initial stage combustion        stability, due to more repeatable and reliable ignition    -   with fixed timing for pilot start of injection (PSOI), methane        emissions decrease with advancing timing for gaseous fuel start        of injection (GSOI) due to more complete mixing    -   with fixed GSOI, methane emissions increase with retarded PSOI,        due to incomplete combustion    -   higher diesel flow rates also help reduce methane emissions    -   retarded combustion timing results in lower NO_(x) emissions,        due to lower combustion temperature    -   PM emissions were significantly higher for tests with higher        diesel flow    -   fuel consumption deteriorates with retarded combustion timing        (that is, retarded PSOI).

Each of the aforementioned U.S. and foreign patent documents andnon-patent publications referred to in this specification are herebyincorporated by reference herein in their entirety.

While particular elements, embodiments and applications of the presentinvention have been shown and described, it will be understood, that theinvention is not limited thereto since modifications can be made bythose skilled in the art without departing from the scope of the presentdisclosure, particularly in light of the foregoing teachings.

1. A safety brake system for a trailer comprising: a brake endcomprising a brake unit adapted to be secured to a trailer andpositioned to brake the trailer, the brake unit being biased towardbraking the trailer; an actuation unit comprising: a mount adapted to besecured to a front end of the trailer; a brake interface operativelysupported by the mount for displacement with respect to the trailer, thebrake interface being operatively connected to the brake unit fordeactivating the brake unit, the brake interface being displaceabletoward a deactivation state in which the brake interface releases thebrake unit from braking the trailer; a probe operatively supported bythe mount for displacement with respect to the trailer, the probe beingdisplaceable between a hitching state in which the probe is adapted tocontact the hitch of a vehicle, and a blocking state in which the probeblocks a hitch coupler of the trailer to prevent hitching of the trailerto a vehicle; a biasing unit to bias the probe against the hitch of thetrailer in the hitching state and toward the blocking state, such thatthe brake unit is actuated when the hitch is separated from the hitchcoupler; and a mechanism operatively connecting the brake interface tothe probe, the mechanism retaining the brake interface in thedeactivated state when the probe is in the hitching state, the mechanismreleasing the brake interface from the deactivated state when the probemoves to the blocking state.
 2. The safety brake system according toclaim 1, wherein the brake interface comprises a lever pivotallyconnected to the mount and manually displaceable to the deactivatingstate.
 3. The safety brake system according to claim 1, wherein theprobe comprises an arm for being manually displaced away from theblocking state.
 4. The safety brake system according to claim 3,comprising a lock unit releasably immobilizing the arm and the brakeinterface with the brake unit braking the trailer and with the probe inthe blocking state.
 5. The safety brake system according to claim 1,comprising a translational joint between the probe and mount for theprobe to move in translation along the mount.
 6. The safety brake systemaccording to claim 1, wherein the probe has a main body operativelysupported on the mount rearward of the hitch coupler, and a probe endhaving a L-shape and having a portion positioned forward of the hitchcoupler, with a movement from the hitching state to the blocking statebeing in a rearward direction.
 7. The safety brake system according toclaim 6, wherein the probe end has a hook engaging with a flange of thehitch coupler in the blocking state.
 8. The safety brake systemaccording to claim 1, wherein the biasing unit is a spring connectedbetween the probe and the mount.
 9. The safety brake system according toclaim 1, further comprising a cable unit having a housing and wire, thewire axially displaceable in the housing, with the cable unit connectedto the brake interface at a first end and to the brake unit at anotherend to transmit deactivation movement of the actuation end to the brakeend.
 10. The safety brake system according to claim 1, wherein the brakeunit has a brake mount per wheel adapted to be connected to the trailer,and at least one arm pivotally connected to the brake mount, to pivot toa braking contact with a rim of the wheel.
 11. The safety brake systemaccording to claim 10, wherein the brake mount has two of said arms perwheel, with the two arms respectively providing braking contact to therim for an own direction of rotation of the wheel.
 12. The safety brakesystem according to claim 1, wherein the mechanism comprises a detentpivotally mounted to the mount and biased against the probe for movingtherewith, and a latch pivotally mounted to the mount and biased againstthe probe, the latch being in oriented for captive engagement with thebrake interface by the detent when the probe is in the hitching state.13. A method for deactivating a safety brake unit of a trailer,comprising: moving a probe away from a blocking state in which the probeblocks an access to a hitch coupler; coupling the hitch coupler to ahitch; and latching a brake interface into a deactivating state in whichthe brake interface releases the safety brake unit.
 14. The methodaccording to claim 13, further comprising manually moving the brakeinterface to the deactivating state when the trailer is unhitched torelease the safety brake unit to align the hitch and the hitch couplerprior to coupling same.
 15. The method according to claim 13, furthercomprising removing a lock preventing movement of the probe and thebrake interface prior to moving the probe.